A new process of refrigeration has recently been developed for achieving efficient refrigeration below 10 Kelvin, particularly at liquid-helium temperatures. A basic description of a system using such process and describing the operating cycle thereof is set forth in U.S. Pat. No. 5,099,650 issued on Mar. 31, 1992 to J. A. Crunkleton. Additional background information concerning such technique is also described in U.S. Pat. No. 4,862,694 issued on Sep. 3, 1989 to J. A. Crunkleton and J. L. Smith, Jr. The more recently issued patent discloses a method for attaining refrigeration at liquid-helium temperatures using a simple and compact multi-stage system configuration. It is helpful in understanding the invention here to review in some detail below the operation of such prior used process.
In a system described in U.S. Pat. No. 5,099,650 which uses two or more stages, in the warmer stages, i.e., those generally at about 20 K. and above, heat transfer occurs between the fluid and the structural material (referred to as a regenerative heat exchange operation), as well as between fluid flowing in separate input and output cooling channels (referred to as a counterflow operation). Fluid flowing in the output channel originates only from the colder stages, i.e., those generally below 20 K., having a connection (e.g., a valve) between the input and output channels. In the colder stages, where obtaining high heat-exchange effectiveness with conventional regenerative structural materials is difficult, heat transfer occurs primarily by counterflow heat exchange. Thus, the technique achieves high heat-exchange effectiveness over the entire temperature range from room temperature down to liquid-helium temperatures by using counterflow heat exchange almost exclusively in the colder stages and by using a combination of both counterflow and regenerative heat exchange in the upper stages, where the inherent mechanical simplicity of a regenerative heat exchange operation may be exploited with high heat-exchange effectiveness.
One embodiment of the technique discussed therein incorporates heat exchangers and piston-cylinder expanders in an integrated two-stage configuration. In that arrangement, the heat exchanger in the warmer stage undergoes both counterflow and regenerative heat-exchange processes, while the colder stage undergoes primarily a counterflow heat exchange process. One exemplary cycle of operation for a two-stage configuration can be described as follows.
Displacement volumes, alternatively referred to as expansion volumes, at each stage of a two-stage configuration are periodically recompressed to a high pressure by reducing the displacement volume in each stage to substantially zero, or near zero, volume. By opening an inlet valve at the warm (e.g., at or near room temperature) end of an input channel, and by increasing the displacement volumes, further fluid under pressure, as supplied from an external compressor, is caused to flow into the input channel at a first relatively warm temperature (e.g., at or near room temperature). The fluid that has been introduced into the input channel is pre-cooled by regenerative and counterflow cooling as it flows through the input channel to the first stage expansion volume at which region it has been pre-cooled to a second temperature below the first temperature. A further portion of the incoming fluid and a residual fluid portion from the previous cycle continue to flow past the first expansion volume in the input channel to the second stage expansion volume at the cold end of the channel. These latter fluid portions are further pre-cooled primarily by counterflow cooling, as well as by some, though much less, regenerative cooling, as they flow in the input channel to the second expansion volume at a third temperature below the second temperature.
The expansion volume at the first stage, i.e., the "warm" stage, is increased, i.e., expanded, so that the compressed fluid therein is expanded from the high pressure at which it had been pressurized to a substantially lower pressure so as to reduce the temperature of the fluid in or near the "warm" displacement volume to a fourth temperature which is substantially lower than the second temperature, but generally higher than the third temperature.
The displacement volume at the second stage, i.e., the "cold" stage, is increased simultaneously with that of the first stage to form an expanded volume at the second stage so that the compressed fluid therein is expanded from the high pressure at which it had been pressurized to a substantially lower pressure so as to reduce the temperature of the fluid in or near the "cold" displacement volume to a fifth temperature which is lower than the third temperature.
At the end of the expansion stroke (at which time maximum expansion volumes exist), the warm exhaust valve and the cold exhaust valve open, which results in blow down if a pressure difference exists across the valves before opening. Although both exhaust valves are opened at some time during the blow down and the constant-pressure exhaust periods, the valves are not necessarily opened or closed at the same time.
The displacement volume at the warm stage is decreased and the low pressure expanded fluid therein is caused to flow back into the input channel from the first stage displacement volume, toward the inlet end of the input channel and thence outwardly therefrom through a "warm" output valve thereat, a small portion thereof alternatively flowing to the cold stage.
Further, the very-low-temperature, low-pressure, expanded fluid which is used to produce the cold environment at the second stage is caused to flow from the "cold" displacement volume, as a result of the decrease in such displacement volume, into an output channel via a "cold" valve and a surge volume thereat, a small portion thereof alternatively flowing through the input channel to the warm stage. The very-low-temperature expanded fluid, which may be two phase, for example, is used to produce a cold environment for a heat load applied thereto, heat being transferred from the environmental heat load to the expanded fluid thereby boiling the two-phase fluid and/or warming the gaseous fluid and cooling the environment. A further heat load may be applied to the warm stage for cooling thereof also.
The lower-pressure fluid, which is caused to flow over a first time duration from the "warm" first-stage displacement volume at the fourth temperature towards the inlet end of the input channel and through the warm exhaust valve thereat, is in intimate contact with the warmer surfaces of the piston and cylinder and exchanges heat with these warmer surfaces thereby warming the fluid exiting from the warm exhaust valve and cooling the piston and cylinder in preparation for the following cycle. This type of heat exchange is commonly referred to as regenerative heat exchange. Simultaneously with such operation, but over a second longer time duration, the expanded low-temperature, low-pressure fluid from the "cold" displacement volume is caused to flow in the output channel at a substantially constant flow rate and at a substantially constant pressure to a fluid exhaust exit at the warm output end of the output channel. During operation, direct counterflow heat exchange is provided between the input and output channels to produce a pre-cooling of incoming fluid in the input channel and a warming of the fluid in the outlet channel to a temperature at or near the first temperature, less allowance of a heat exchange temperature difference prior to its exit therefrom. The warm exiting fluid from both the input and output channels is compressed, as by being supplied to an external compressor system, so as to supply fluid under pressure from the compressor system for the next operating cycle.
Residual portions of the expanded fluid which resulted from the expanded operation of a previous cycle remain in the displacement volumes and in the input channel. Such remaining fluid may undergo recompression if the warm and cold exhaust valves are closed before minimum displacement volumes are reached. The device is now ready to execute the next expansion cycle. The compressed fluid from the compressor system is next supplied via the input channel to the first and second stage displacement volumes. The fluid flowing to the first stage displacement volume is pre-cooled by regenerative heat exchange with the piston and cylinder structures, and by counterflow cooling by the cold fluid flowing in the output channel. The fluid flowing to the second stage displacement volume is primarily pre-cooled by counterflow heat exchange with the cold fluid flowing in the output channel, although there may be some, but much less, pre-cooling due to regenerative cooling.
The overall compression, intake, expansion, and exhaust process is then repeated, the fluid in the displacement volumes and in the input channel being again periodically compressed and the expansion thereof occurring as before.
The size of the heat load (i.e., including both an applied heat load and/or parasitic heat leaks) at either stage has a relatively large impact on the type of heat exchange operation at the warm stage. If the heat load at the cold stage is much smaller than that at the warm stage, regenerative heat exchange dominates at the warm stage. If the heat load at the cold stage is relatively larger than that at the warm stage, counterflow cooling may account for most of the heat exchange at the warm stage. This is because a relatively larger heat load on the cold stage requires more mass flow to the cold stage. This larger mass flow rate returns to the compressor primarily through the output passage, which results in more counterflow heat exchange in the warm stage.
The power requirement of the refrigerant apparatus can be decreased by increasing the number of precooling stages. For example, two precooling stages, both operating above 20 K., significantly reduce the power requirement. In this example, both precooling stages operating above 20 K. use a combination of both counterflow and regenerative heat exchange.
Preferred configurations of this refrigeration method prescribe annular passages between concentric tubes to be the input and output channels. The input channel is formed by the gap between the piston and cylinder and the output channel is formed by the gap between the cylinder and an outer shell that surrounds the cylinder. If a hollow piston is used, the volume inside the piston is at vacuum to reduce the heat leak. In this arrangement, gap nonuniformities can lead to flow maldistributions in the channels which result in reduced heat exchanger performance. For example, if the piston or cylinder is not perfectly straight and round, or if the piston is not perfectly centered inside the cylinder by some centering means, then the piston-to-cylinder gap is not constant along the circumference at all locations along the length, which results in flow maldistribution. A spiral passage is constructed between the cylinder and outer shell to direct the output-channel flow around the cylinder to reduce the effect of flow maldistribution in the piston-to-cylinder gap. An object of the present invention is to further reduce or substantially eliminate flow maldistributions that are present in the previous systems.
Also in this concentric tube arrangement, the use of the earlier described annular gaps, where the input flow travels axially between the piston and cylinder or where the output flow spirals between the cylinder and outer shell, allows transitions between laminar and turbulent flow conditions over a wide range of temperatures during various parts of the cycle. Heat exchanger performance varies considerably depending on whether the flow is laminar or turbulent, which makes design of the heat exchangers difficult. Another object of the present invention then is to provide a heat exchanger configuration which causes a continuous mixing of the flow in the heat exchangers as the flow proceeds along its length, so the flow is never fully developed into a laminar or turbulent regime but rather is continually mixed.
Different types of drive mechanisms to reciprocate the piston are presented in the aforesaid Crunkleton patent for specific cryocooler embodiments. Two drive mechanisms disclosed therein are described again here for illustrative purposes.
One type of drive mechanism in which no energy is stored uses a pressure-balanced piston, meaning that, in the ideal case, the pressure is equal on all piston surfaces so that no net force is placed on the piston. In an actual device, the pressure is approximately equal at each cross section along the axis of the piston; however, due to pressure drops in the precooling heat exchanger which cause an axial end-to-end pressure difference, a net axial force results on the piston. Because of the phasing of this pressure difference during pressurization and depressurization from the warm end, the resulting axial force can be used to reciprocate the piston. Subtracted from this axial force is the drag of any piston seals as well as other frictional forces resulting from piston motion. Another common pressure-balanced-piston drive mechanism is reciprocated using a stepper motor in combination with a scotch-yoke mechanism. Either configuration would be well known to those in the art.
Another type of drive mechanism, in which energy is stored for use later in the cycle, uses a piston that is not pressure balanced. The resulting force and piston displacement yield an external work transfer from the cold working volume to the room-temperature end. This work can be temporarily stored and then later used for recompression and to overcome any friction such as due to sliding bearings and seals. A typical mechanical configuration to achieve this operation employs a flywheel for energy storage, which would be well known to those in the art.
While the system described in the aforesaid Crunkleton patent provides refrigeration, no technique is disclosed therein to customize each stage of refrigeration to improve performance and to provide high reliability. The present invention consists of several improvements that are intended to increase performance and reliability of components operating over the entire range of temperatures from room temperature down to liquid-helium temperatures. In particular, various improvements have been made to the precooling heat exchangers, the load heat exchangers, mechanisms to control flow in the precooling heat exchangers, mechanisms to control flow to and from the cold head.